Automotive vehicles have used drum brake assemblies on either the front or the rear axles, or both, for many years. In recent years it has also become commo to equip automotive vehicles with disc front brake assemblies and drum rear brake assemblies.
The type of drum brake assembly commonly known as a duo-servo drum brake has been used since the 1920's on millions of automotive vehicles including passenger cars, light trucks, heavy duty trucks, buses and trailers. Duo-servo drum brake assemblies are characterized by having a primary brake shoe and a secondary brake shoe with one set of adjacent shoe ends being acted upon by a wheel cylinder or cam to expand the brake shoes for braking engagement with a brake drum, and the other set of adjacent shoe ends typically being connected through an adjuster strut or link mechanism for maintaining the brake shoes in adjustment as the brake linings wear. The shoes are mounted on a backing plate and are capable of limited sliding movement in a somewhat arcuate direction as well as expanding outward to engage the friction surface of the brake drum. Other types of brake shoe assemblies have also been used at various times. These include leading-trailing brake shoe assemblies and leading-leading brake shoe assemblies, by way of example. However, the duo-servo type of brake shoe assembly has been most commonly used, particularly in passenger and light truck automotive vehicles.
It is a characteristic of the duo-servo drum brake that the hydraulic servo action obtained by pressurizing the wheel cylinder is augmented by a mechanical servo action of the brake shoes. In normal forward rotation of the brake drum, the primary brake shoe engages the rotating drum. The rotation of the drum with the shoe in engagement therewith arcuately drives the shoe mechanically toward the adjusting strut, generating additional brake actuating force transmitted through the adjusting strut and applied to the secondary brake shoe. Thus, the hydraulic servo action and the mechanical servo action cooperate to increase the brake torque at a higher rate than the gain without mechanical servo action.
At times it has been desirable to reduce the output variation of drum brakes which is a well-known condition that has existed for many years. Efforts to accomplish this have usually focused upon modifying both the shape and location of the anchor pin. It has been common in recent years, particularly with front wheel drive vehicles having disc brakes on the front wheels, to employ a wide, oval shaped anchor pin. This anchor pin reduces the gain of the foundation brake by permitting the anchor pin to be mounted closer to the friction surface of the drum, thereby reducing the leverage of the friction force reaction at the anchor. Another approach has been to locate the primary brake shoe lining below the center of the shoe, which generates a tendency to reduce the mechanical servo action of the primary brake shoe on the secondary brake shoe, thus reducing the gain of the duo-servo brake design.
Studies of vehicles employing leading-trailing drum brakes and duo-servo drum brakes have shown that neither type of drum brake is completely stable throughout the brake useful life in normal service. Drum brakes tend to increase in specific torque with usage. This trend in duo-servo brakes is somewhat offset by the tendency for the hold-off pressure to increase. Hold-off pressure is the hydraulic pressure applied to the wheel cylinder required to overcome the shoe retractor springs, the drag forces on the shoe exerted between the shoe and the backing plate, and wheel cylinder efficiency losses. These factors tend to increase with usage.
Various approaches to accommodate the brake system to output variations include the use of lower gain brake designs such as the disc brake or the leading-trailing drum brake. Disc brakes are known to decrease their brake effectiveness somewhat with use. Existing leading-trailing brakes have been known to increase brake output, decrease brake output, or remain relatively unchanged in brake output during their service life. This does not address the issue of variation in brake output during the useful life of the brakes, but instead attempts to somewhat reduce the effect of such variations through the us of such lower gain designs. This requires major retooling and validation, adding to the cost.
Contact pressure distribution between the friction material forming the brake lining and the drum brake friction surface results from a complex relationship between the anchor pin location, the wheel cylinder location, the magnitude of the hydraulic brake actuating pressure applied, the primary and secondary shoe and lining stiffness, and the drum stiffness and deflection during loading. This contact pressure distribution may also be affected by the work history of the brake as the brake lining wears.
The history of analysis, theory and understanding regarding drum brakes has indicated that such analysis and understanding has not kept pace with the practical design and production of drum brakes. A reasonably comprehensive listing of literature and source documents relating to such analysis, theory and understanding is as follows:
(1) Acres, "Some Problems in the Design of Braking Systems", The Journal of the Institution of Automobile Engineers, London, Vol. XV, No. 1, Oct. 1946, pp. 19-49. PA0 (2) Fazekas, "Some Basic Properties of Shoe Brakes", Journal of Applied Mechanics, published by the American Society of Mechanical Engineers, Vol. 25, March 1958, pp. 7-10. PA0 (3) Robinson, "Brake Design Considerations--Some Notes on the Calculation of Shoe Factor", Automobile Engineer, September 1959, pp. 340-348. PA0 (4) Oldershaw and Prestidge, "Brake Design Considerations--An Approach Based on the Concept of a Centre of Pressure for the Reaction Between the Shoe and the Drum", Automobile Engineer, April 1960, pp. 157-159. PA0 (5) Steeds, "Brake Geometry--Theory of Expanding Rigid Types", Automobile Engineer, June 1960, pp. 261-262. PA0 (6) Newcomb, "Determination of the Area of Friction Surfaces of Automotive Vehicles", Journal of Mechanical Engineering Science, Vol. 2, No. 4, December 1960, pp. 312-324. PA0 (7) Ferodo Ltd., "Friction Lining Calculations--The Sizing of Brake and Clutch Facings", Automotive Design Engineering, Vol. 4, July 1965, pp. 93-94. PA0 (8) "Motor Vehicle Performance--Measurement and Prediction", Section 6.0, "Braking Systems and Braking Performance", University of Michigan Engineering Summer Conferences, Conference Proceedings, July 9-13, 1972, pp. 157-208. PA0 (9) Millner and Parsons, "Effect of Contact Geometry and Elastic Deformations on the Torque Characteristics of a Drum Brake", The Institution of Mechanical Engineers Proceedings 1973, Vol. 187, pp. 317-331, and Discussion, pp. D105-108. PA0 (10) "Engineering Design Handbook--Analysis and Design of Automotive Brake Systems", DARCOM Pamphlet DARCOM-P 706-358, HQ, U.S. Army Materiel Development and Readiness Command, December 1976, Chapters 1, 2 and 14. PA0 (11) Orthwein, "Estimating Torque and Lining Pressure for Bendix-Type Drum Brakes", SAE Paper No. 841234, October 1984. PA0 (12) USG 2456 dated January 13, 1986, entitled "General Motors Corporation Response to Proposed FMYSS 135--Passenger Car Brake System, Docket No. 85-06, Notice 1, Appendices Nos. 1 (Glossary of Terms) and 12 (variability). This report is a public record filed with the Associate Administrator for Rule Making, NHTSA, 400 Seventh St. SW, Washington, D.C. 20590.
In various ones of these documents one may find analyses of many different drum brake configurations including leading-leading, trailing-trailing, leading-trailing, and duo-servo brakes. Document (12) above provides an excellent glossary of brake terms in its Appendix 1 and a thorough discussion of brake variability in its Appendix 12. Early problems which the analysis of drum brake designs tried to solve included developing sufficiently powerful designs and predicting the output of such brakes prior to the construction of test parts.
The development of the duo-servo drum brake, where the primary brake shoe is used to increase the applied load to the secondary brake shoe to increase the torque output of the drum brake, yielded a very powerful brake in a relatively small physical package which did not require large displacement hydraulic systems. These advantages are still found in this type of drum brake as it is used today. The early modeling efforts assumed a graphical approach to calculate "shoe factors", a phrase meaning the specific torque attributable to each individual brake shoe. These early efforts dealt primarily with the geometrical leverage ratios associated with wheel cylinder location, anchor pin location, etc., and assumed a uniform pressure distribution along the friction interface. It appears that the uniform pressure distribution was assumed because such an assumption was then convenient for analytical solution of the brake design. Examples are found in Acres, Fazekas, Steeds and Oldershaw noted above, as well as other treatises. Later design and theory efforts recognized that the uniform distribution assumption did not correctly predict the measured specific torque of drum brakes. This led to analysis based on the assumption of a sinusoidal pressure distribution.
All brake design efforts prior to the early 1970's assumed that both the drum and the brake shoe as well as the lining on the shoe were rigid and did not undergo elastic deformation under load even though the theorists realized that this was not the case. This assumption was made in order to simplify the analytical efforts. It was only with the advent of high speed digital mainframe computers and the development of finite element analysis that the shoe and drum deflection have been able to receive any theoretical effort. The 1973 article by Millner and Parsons, noted above, dealt with these deflections. That article also contains the observation that the pressure distribution may change due to work history of the brake. Heel-toe contact was analyzed for a leading-leading drum brake and shown to cause an increase in brake output. That article and the discussions appended thereto also discussed variations with changes in coefficients of friction of different lining materials. Millner and Parsons touched on the effect of lining arc length in relation to the "shoe factor", but had no discussion or other indication which would point to the desirability of short lining arc lengths as now proposed.
Historically, brake system engineers have held a skeptical view point with regard to the ability of drum brake mathematical models to accurately predict the specific torque characteristics of practical drum brake designs. This has seemingly resulted from their experience in seeking to accurately employ the model predictions in designing vehicle brake systems. For example, minute changes in the dimensions of the component parts of the typical drum brakes in use for many years, or changes in lining composition and thus the friction coefficient of the lining, may result in changes in drum brake specific torque, and thus similar changes in the torque output of a brake. A missing piece of the puzzle has always been the knowledge of the actual lining-drum friction interface pressure distribution. The classical assumptions of rigidity and consistent pressure distributions have been coupled with the practical experience of the brake system engineers to produce a significant body of skeptics regarding the usefulness of models and drum brake analysis.
Several of the publications noted above touched on the arc length of the brake lining material. There is a common thread of thought throughout them that there should be as much arc length reasonably practical, considering the need for cooling. The unit load on the friction material, as well as its deformation and wear characteristics, further led to the broad adoption in duo-servo brakes of primary and secondary shoes with arc lining lengths in the ranges of 90.degree. to 100.degree. and 110.degree. to 120.degree. , respectively, and particularly respective lining arc lengths of about 97.degree. and 117.degree..
Acres proposed the use of four shoes with 75.degree. to 80.degree. arc length each, but found that there was insufficient heat dissipation, probably because the linings covered most of the drum friction surface at all times and hindered cooling. He expressed a thought that maybe he should have tried only two such shoes, with double duty on the lining, just as a check, but did not pursue this further. Instead, Acres cited with approval an article by Super in the May 1946 issue of the SAE Journal that the total arc length should be reduced to 240.degree. because of the heat dissipation problem. Assuming an even division between two shoes, each shoe lining would then have an arc length of 120.degree.. Acres also discussed the effects of various friction lining materials, noting the changes which occur with them. He made it obvious that different materials yielded very different results.
The Steeds paper made assumptions relating to rigidity and full contact of the lining with the brake drum along its whole lining arc. Steeds gives no indication that he was considering lining arc lengths other than those in common use at the time and about the same as the other prior art.
Oldershaw and Prestidge, contemporaries of Steeds, touched on lining arc lengths, and concluded that the limits of lining arc length for a shoe were 80.degree. -135.degree., with graphs clearly showing their position that shoe lining arc lengths of 100.degree. -120.degree. were the most practical.
Some typical patent disclosures discussed below indicate various developments in brake shoes and linings which generally relate to the problems solved by the use of the invention herein disclosed and claimed.
U.S. Pat. No. 2,750,006, entitled "Brake" and issued June 12, 1956, dealt with improvements in brake lining wear and chatter. The invention was that of a brake lining of a pivoted shoe assembly in which the brake lining has a cylindrical curvature at least at both ends greater than the cylindrical curvature of the internal surface of the brake drum with which the shoe is brought into contact during rotation of the wheel being braked.
U.S. Pat. No. 2,818,941, entitled "Motor Vehicle Brake Construction" and issued Jan. 7, 1958, proposed that the primary lining of a duo-servo brake assembly be made considerably narrower in width than the secondary lining. By proper proportioning of these widths, substantially uniform unit pressures between each of the linings and the drum was considered to be obtained. It was felt that this would reduce brake fade under severe braking conditions.
U.S. Pat. No. 2,848,073, entitled "Friction Device Such As A Brake" and issued Aug. 19, 1958, proposed the use of a plurality of short lining elements on each brake shoe. The lining elements on both shoes together occupied a total of about 60.degree. of the total circumference of 360.degree.. Thus each brake shoe has, as shown and described, about a total of 30.degree. of arc length, divided into two sections apparently of about 15.degree. each, for each shoe. The inventor recognized that additional cooling area contributed to the satisfactory performance of his brake. However, he had no recognition of the pressure distribution or the deflection characteristics of shoes and drums. He also did not recognize any advantages with regard to the different coefficients of friction of different materials, even though he suggested the use of non-metallic or composition aligning elements. He arranged these shoes in a leading-trailing shoe assembly. The inventor limited his friction devices to having a plurality of short pads on each shoe, the pads being spaced from the horizontal axis of the shoe and therefore more adjacent each of the shoe ends than the lining herein disclosed and claimed.
U.S. Pat. No. 2,910,145, entitled "Brake Structure" and issued Oct. 27, 1959; and U.S. Pat. No. 3,013,637, entitled "Brake" and issued Dec. 19, 1961, show other configurations using small multiple friction lining pad units in a duo-servo brake context. In a similar fashion, U.S. Pat. No. 3,007,549, entitled "Friction Controlling Means" and issued Nov. 7, 1961, provided gaps between various friction lining sectors on the secondary brake shoe of a duo-servo brake arrangement, or on both brake shoes of a leading-trailing brake arrangement.
U.S. Pat. No. 3,029,901, entitled "Vehicle Drum Brake" and issued April 17, 1962, concerns a duo-servo drum brake with a modified anchor arrangement to limit the self-energizing force on the secondary shoe. The disclosure also shows two sectors of brake lining materials provided on the primary shoe and the secondary shoe, the sectors being mounted at the heel and toe of each shoe and considerably spaced apart and well away from the horizontal axis of the assembly. It is noted that the patentee utilized the brake analysis of Acres as set forth in the article referred to above.
The technology of asbestos-containing friction materials for drum brakes has evolved over a period of more than 70 years with improvements in wear, fade and friction stability continuing over this time span. Some of the above noted articles and papers touch on this. The technological development of non-asbestos friction materials, particularly for drum brake applications, has been occurring to any major extent only in about the last 10 years. While the non-asbestos technology has been successful in the development of metal fiber reinforced materials for disc brake applications, attempts to use a wide range of mineral, metal and man-made fibers in drum brake applications has met with only limited success. Although some limited activity occurred 25 to 30 years ago, the results were never widely adopted, at least in part because of wear and noise problems. Among the issues faced in drum brake applications were and are wear, friction, braking effort variations, and noise performance of non-asbestos materials.
Because it is desirable to eliminate the use of asbestos products, and because the more suitable non-asbestos friction materials generally have higher friction coefficients than the presently used asbestos friction materials, the brake industry has tended to plan to meet a proposed government ban on asbestos by designing future vehicles with lower gain rear brake systems such as leading-trailing drum brakes. To produce equivalent brake torque, these alternative brake systems require larger hydraulic displacement, which has a design impact on master cylinder size and brake booster response as well as brake pedal ratio. The brake engineer also has to consider the desired stopping distances and the effects of such changes thereon, as well as modulation by either manual or anti-lock brake control systems, particularly when displacements are increased. Even if the various factors so noted can be treated through innovative system designs, the concern of variability in brake output with higher friction coefficient materials must be addressed to insure acceptable brake performance throughout the reasonable useful life of the brakes on the vehicle. If the use of asbestos products are prohibited in brake linings manufactured for service use, an issue that is also in need of consideration is that of servicing existing vehicles with non-asbestos lining substitutes when the brakes were designed for use with asbestos friction materials.